Rotary slant shaft type gas compressor with multi-stepped exhaust system

ABSTRACT

A rotary slant shaft type gas compressor having a multi-stepped exhaust system is provided, which includes: a driving shaft fixed with a cylinder head formed with gas holes; a gas guide member for intake of gas and discharge of compressed gas; a case head member coupled with the driving shaft and formed with an intake port and three exhaust ports; a valve plate member fixed on an inner surface of the case head member to contact an outer surface of the cylinder head, and formed with a gas intake valve groove and three gas exhaust valve grooves; a cylinder block formed with cylinder bores, integrally coupled with the cylinder head, and slidably inserted by pistons; and a swivel plate member connected to the cylinder block and the pistons and converting the rotation force to reciprocation motion.

BACKGROUND OF THE INVENTION

(a) Field of the Invention

The present invention relates to gas compressors, and more particularlyto a rotary slant shaft type gas compressor having a multi-steppedexhaust system for selectively exhausting gas compressed in a cylinderaccording to a pressure of an exhaust channel.

(b) Description of the Related Art

A compressor is a machine for increasing a pressure and a potentialspeed of a medium by applying power from the outside. Such compressorsare called fluid compressors since a fluid is an object of thecompressor regardless of the state of the medium being compressed. Asthe media which may be compressed by the compressor, there are gassessuch as air, nitrogen, oxygen and the like, and liquids such as oils orrefrigerants. Even though a compressor to be described hereinafter maybe used for compressing liquids such as oil, a gas compressor thatcompresses gasses such as air will be principally described.

As a publicly known gas compressor, there is a reciprocating compressorthat compresses gas with a piston that carries out a simplereciprocation motion.

In general, the reciprocating compressor is formed with a cylinder, apiston reciprocating in the cylinder, and a cylinder head comprising anintake valve and an exhaust valve at an end of the cylinder, like anengine of a vehicle. In such a reciprocating compressor, intake,compression and exhaust of gasses are carried out while opening andclosing the intake valve and the exhaust valve according to a gaspressure in the cylinder as the piston rectilinearly reciprocates in thecylinder.

This reciprocating compressor has, however, a disadvantage in that theintake valve and the exhaust valve mounted in the cylinder head directlycontact the cylinder head or the piston during the gas compressionstroke. The collision of the valves primarily induces mechanical noise,and bending or damage of the valves occurs in long-term use. Further,the reciprocating compressor has disadvantages in that a pulsationphenomenon is generated in the case of gas compression since the intakeand the exhaust of gas occurs alternately in the cylinder, and thatfriction noise is generated by the instant expansion of the gas whenopening or closing the valves.

An intake/exhaust muffler is provided to resolve the noise problem ofthe reciprocating compressor. However, if a muffler is mounted on thereciprocating compressor, the compressor itself becomes complicatedmechanically and the number of required parts increases. Further, thegas resistance is increased due to the mounting of the muffler, therebydegrading performance of the compressor.

A slant shaft type compressor is disclosed as another gas compressor inJapanese Laying-open Publication No. 61-65081 (Apr. 3, 1986).

In the compressor disclosed in the publication No. 61-65081, rotationforce of a rotation shaft is transmitted to a swivel plate, which isconnected to pistons, for converting the rotation motion to arectilinear reciprocation motion. In the compressor, a cylinder blockformed with six cylinders is fixed to the rotation shaft and respectivecylinders in the cylinder block are formed in a structure such that asurface facing a piston is open. The open cylinder is closed by a floatvalve formed with an intake/exhaust hole and a compressor case headcontacts a rear surface of the float valve. A rubber ring is interposedbetween the float valve and the case head for preventing leakage of gascompressed in the respective cylinders.

In this compressor, if a driving shaft is rotated by rotation forcetransmitted from an external power supply, the cylinder block fixed tothe driving shaft rotates together with the driving shaft, and theswivel plate connected to an end of the driving shaft rotates inresponse to the rotation of the driving shaft, so that the respectivepistons rectilinearly reciprocate in the respective cylinders, insequence.

According to the characteristics of this compressor, the respectivecylinders rotate as being opened while the float valve and the case headdo not move. The respective cylinders take in the gas through the intakehole of the float valve for gradually compressing the gas whilerotating, and exhaust the compressed gas through the exhaust hole of thefloat valve toward a gas channel formed in the case head. In the abovecompression stroke, the float valve moves close to the cylinder block bya difference of gas pressures applied to a sectional area of thecylinder and a sectional area of the valve.

Comparing the compressor disclosed in the publication No. 61-65081 withthe prior art reciprocating compressor, the piston of the compressor of61-60851 reciprocates in parallel with the driving shaft direction,thereby allowing the manufacture of the compressor to be compact.Further, the compressor of 61-65081 does not employ reciprocatingintake/exhaust valves but a fixed float valve, so that the mechanicalnoise caused by the direct collision between the valves and the cylinderhead may be completely prevented. Furthermore, the compressor of61-65081 exhibits compression efficiency and noise characteristics dueto the gas pressure difference equal to the prior art reciprocatingcompressor in the case of continuous operation under a rated load.

In spite of the advantages described above, the compressor of 61-656081has a serious disadvantage in that the cylinder block has to rub thefloat valve to maintain the seal between the rotating cylinder block andthe stationary float valve, thereby causing abrasion of parts due to thecontinuous friction therebetween. In order to remove friction heatgenerated by the friction, the gas to be compressed has to belubricative. Therefore, the gasses compressed in the compressor arelimited to those having the lubrication property.

Further, the compressor has a disadvantage in that additional parts foremitting heat to the inside or the outside or absorbing the heat isneeded, since the compression heat generated in the process of thecompression of the gas in addition to the friction heat is very high.However, the compressor of 61-65081 does not suggest any heat removalparts, so durability of the compressor is degraded and gas compressionefficiency is decreased by the various heat generated in actual use.

Considering the compressor of 61-65081 aerodynamically in view of thestructure of the compressor, the compressor has a very big differencebetween a maximum pressure (Pm) in a compression section and an exhaustpressure (Pd) in an exhaust section. In this case, as the pressuredifference between the two sections becomes larger, the aerodynamicnoise generated when compression gas of a high pressure is discharged toa low pressure state becomes larger. Considering the compressor of61-65081 with the prior art compressor on this issue, the compressor of61-65081 exhibits a larger aerodynamic noise than the prior artreciprocating compressor due to such a big pressure difference.

Considering a compression load in the cylinder generated duringoperation, the compressor of 61-65081 exhibits a change width of thecompression load per a unit time period much larger than that of theprior art reciprocating compressor. As the change of the compressionload in the cylinder becomes larger, an axial force load applied to thedriving shaft becomes larger. Therefore, in the compressor of 61-65081,the axial force load which is proportional to the compression load isapplied to the swivel plate connected to the end of the driving shaft,directly influencing ball bearing parts mounted between a lower part ofthe swivel plate and the case, thereby degrading the durability of thecompressor itself.

As described above, the compressor of 61-65081 has problems caused bythe structure in spite of the various advantages over the prior artreciprocating compressor, so the compressor has a commercial limitationas a gas compressor.

Therefore, the demands for a new compressor of a structure that maymaintain the basic characteristics of the slant shaft type gascompressor but resolves the disadvantages of the compressor of 61-65081to minimize the aerodynamic noise, improve the durability of parts andaccessories, increase the energy efficiency, minimize the number ofrequired parts, and achieve loadless operation are increased togetherwith demands for diversifying of the gasses to compress.

SUMMARY OF THE INVENTION

The present invention is derived to resolve the above problems of theprior art, and it has an object to provide a rotary slant shaft type gascompressor for discharging gas compressed in cylinder bores, not atonce, but selectively in association with an external pressure.

It is another object of the present invention to provide a rotary slantshaft type gas compressor with a structure that may be designedaerodynamically for minimizing the noise mechanically andaerodynamically.

It is a further object of the present invention to provide a rotaryslant shaft type gas compressor in which power required for compressinggas may be minimized to maximize the energy efficiency.

It is a still another object of the present invention to provide arotary slant shaft type gas compressor in which a change of acompression load per unit time period may be minimized for improving thedurability.

It is a still further object of the present invention to provide arotary slant shaft type gas compressor in which gas to be taken intorespective cylinder bores is first circulated through a crank chamberand then introduced into the cylinder bores.

It is a still another object of the present invention to provide arotary slant shaft type gas compressor capable of operating loadlesslywith a high efficiency.

It is a still another object of the present invention to provide arotary slant shaft type gas compressor in which friction heat generatedinside and compression heat generated by air compression may beeffectively emitted.

In order to achieve the above objects of the present invention, a rotaryslant shaft type gas compressor includes a valve plate contacting arotating cylinder head and formed with an intake groove and a pluralityof exhaust grooves, wherein the valve plate is fixed to a case head forselectively discharging gas compressed in cylinder bores.

In more detail, the rotary slant shaft type gas compressor includes: adriving shaft integrally formed with a cylinder head perpendicular to adriving shaft axis, the cylinder head being formed with a plurality ofgas holes on a concentric circle at uniform intervals; a gas guidemember formed with an intake manifold for intake of gas from the outsideand an exhaust manifold for discharging gas compressed in cylinder boresto the outside; a case head member for rotatably supporting the drivingshaft formed with at least one intake port for supplying the gas takenin through the intake manifold to the inside of the cylinder bores, andtwo or more exhaust ports for discharging the gas compressed in thecylinder bores to the exhaust manifold; a valve plate member fixed on aninner surface of the case head member to contact an outer surface of thecylinder head, and formed with a gas intake valve groove and at leasttwo gas exhaust valve grooves on a periphery on which the gas holesmove, the gas intake valve groove supplying the gas taken in through theintake port to the inside of the cylinder bores and the gas exhaustvalve grooves discharging the gas compressed in the cylinder bores tothe exhaust ports; a cylinder block formed with a plurality of cylinderbores in parallel with the driving shaft and having a surface integrallycoupled with the cylinder head, and having an opposite surface slidablyinserted with pistons in respective cylinder bores for compressing theintake gas in the respective cylinder bores; a swivel plate memberconnected to a center part of the cylinder block with a coupling, andconnected to the plurality of pistons via piston rods, for convertingthe rotation force transmitted from the driving shaft to rectilinearreciprocation motion to be transmitted to the pistons; a case end plateformed with a slant surface for supporting the swivel plate member; anda case coupled with the case head member and the case end plate forincorporating the cylinder block and the swivel plate member.

In the rotary slant shaft type gas compressor of the present invention,the respective exhaust ports of the case head member incorporaterespective check valves for selectively discharging the compressed gasvia the respective exhaust grooves of the valve plate member accordingto an internal pressure of a compression tank.

The case head member and the driving shaft are formed with a circulationcircuit for introducing the gas introduced from the intake manifold tothe cylinder bores via a sealed crank chamber formed inside the case, sothat aerodynamic noise possibly generated when compressed air remainingin the cylinder bores after an exhaust stroke is finished is met to newintake gas may be limited in the case, thereby minimizing the noise.

Further, a tension ring and a ring-shaped plate spring are insertedbetween an inner surface of the case head member and the valve platemember, so that the gap possibly generated by the friction between thevalve plate and the cylinder head for a long term use may be completelyprevented.

The rotating cylinder block and the swivel plate member are connected bya universal coupling or a spring coupling, so that the mechanical noisegenerated while the operation of the compressor may be minimized.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of a rotary slant shaft type gascompressor according to the present invention;

FIG. 2 is a perspective view of a gas guide member of the gas compressoraccording to the present invention;

FIG. 3A is a perspective view of a case head of the gas compressoraccording to the present invention;

FIG. 3B is a cross-sectional view of the case head taken along the lineI—I of FIG. 3A;

FIG. 3C is a cross-sectional view of the case head taken along the lineII—II of FIG. 3A;

FIG. 4 is a perspective view of a tension ring of the gas compressoraccording to the present invention;

FIG. 5 is a perspective view of a valve plate of the gas compressoraccording to the present invention;

FIG. 6A is a cross-sectional view of a rotary slant shaft type gascompressor according to another preferred embodiment of the presentinvention;

FIG. 6B is a cross-sectional view of the spring coupling of the gascompressor taken along line III—III of FIG. 6A;

FIG. 6C is a cross-sectional view of a spring coupling according toanother preferred embodiment of the present invention;

FIG. 7 is a cross-sectional view of a rotary slant shaft type gascompressor according to another preferred embodiment of the presentinvention;

FIG. 8 is the cross-sectional view of FIG. 1, showing gas intake andexhaust process of the gas compressor;

FIG. 9A is a view for explaining the gas intake, compression and exhauststrokes in the valve plate of the gas compressor according to thepresent invention;

FIG. 9B is a view for explaining the gas compression characteristicswhen cylinders rotate one cycle in the valve plate of the gas compressoraccording to the present invention;

FIG. 10A is a view for explaining the gas compression characteristicswhen the gas compressor is operating;

FIG. 10B is a view for explaining gas compression characteristics when aprior art reciprocation type compressor is operating; and

FIG. 10C is a view for explaining gas compression characteristics when aprior art slant shaft type gas compressor is operating.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

Reference will now be made in detail to preferred embodiments andmodifications of the present invention, examples of which areillustrated in the accompanying drawings.

As shown in FIG. 1, main parts of a rotary slant shaft type gascompressor according to the present invention are housed in acylindrical case 1. The case 1 is fixed with a case head 30 and a caseend plate 3 at opposite side surfaces by bolts, and rubber rings 4 areinserted into each coupling surface of case parts 1, 30 and 3, so that acrank chamber 70 in the case is sealed from the outside.

The gas compressor received in the case 1 includes: a driving shaft 10for transmitting the rotation force supplied from an external powersupply to a swivel plate; a gas guide member 20 for intake and exhaustof gas from or to the outside; a case head member 30 for supplying theintake gas to a cylinder and selectively exhausting compressed gasaccording to a pressure of a compression tank; a valve plate member 50fixed on an inner surface of the case head member for supplying gas intothe rotating cylinder and exhausting the compressed gas; a cylinderblock 60 incorporating a plurality of pistons for compressing the gas;and a swivel plate member 80 for converting the rotation motion of thedriving shaft to rectilinear reciprocating motion.

The driving shaft 10 is extended as a central axis of the case 1 and isrotatably fixed to a boss part of the case head 30 by a ball bearing 11and a taper roller bearing 12.

The driving shaft 10 has an end fixed with a driving pulley 5transmitting the rotation force generated from the external power supply(not shown) to the driving shaft 10, and the other end is integrallyformed with a cylinder head 13 for sealing cylinder bores 61 of thecylinder block 60. The cylinder head 13 is integrated with the drivingshaft and is formed in the shape of a circular disc, and it has six gasholes 14 formed concentrically with the driving shaft at uniformintervals.

The driving shaft 10 is formed with a shaft chamber 15 inside it andwith axial ports 16 perpendicular to the shaft direction for serving asflow channels to supply the gas introduced into the crank chamber 70 tothe cylinder bores 61.

In the driving shaft, an end where the shaft chamber 15 is formed has aliquid introduction preventing shoulder 17, so that dispersedlubrication oil which may flow along a side surface of a block chamber63 of the rotating cylinder block is prevented from being introducedinto the cylinder bores 61.

The gas guide member 20 is formed cylindrically as shown in FIG. 1 andFIG. 2, and is fixed with an intake tube 21 and an exhaust tube 22 onits cylindrical body. The intake tube 21 serves to introduce gas to becompressed into the compressor and is attached with a filter (not shown)at an outside. The exhaust tube 22 serves to discharge the gascompressed in the compressor to the compression tank (not shown) as astorage tank, as it is communicated with the compression tank. Theexhaust tube 22 incorporates a check valve 27 therein for preventing thecompressed gas in the compression tank from flowing back to the insideof the cylinder bores. The gas guide 20 is formed with an intakemanifold 23 communicated with the intake tube 21, and an exhaustmanifold 24 communicated with the exhaust tube 22 on a samecircumference in a bottom surface thereof. The intake tube 21 and theexhaust tube 22 are attached to an auxiliary intake tube 25 and anauxiliary exhaust tube 26 respectively on a side surface, such that theauxiliary intake tube 25 and the auxiliary exhaust tube 26 are connectedto each other for minimizing a load of the compressor without stoppingthe operation thereof when compression of the gas is no longernecessary.

Such an intake compression stroke as alluded to above is called loadlessoperation. In such a loadless operation state, the pressure compressedin the compression tank is prevented from flowing back to the exhausttube 22 by the check valve 27 incorporated in the exhaust tube.Reference number 28 represents bolt holes for closely contacting andfixing the gas guide to the case head 30 by bolts.

The case head 30 is, as shown in FIG. 1 and FIG. 3, formed with heatemission fins 31 and guide grooves 32 for fixing the gas guide member onan outer surface thereof, and mounting grooves 33 for mounting a thrustbearing 18 for uniformly maintaining a gap from the cylinder head 13,plate grooves 34 for fixing the valve plate 50, and a ring groove 35 forinserting a tension ring 40 which contacts the valve plate with thecylinder head on an inner surface thereof. The case head 30 is formedwith a boss part 36 for receiving the ball bearing 11 and the taperroller bearing 12 to support the driving shaft 10 in a center thereof.The heat emission fins 31 formed on the outer surface of the case head30 are arranged radially with respect to the driving shaft, so thatcooling air flows smoothly to the outside by way of an external air fan(not shown). In order to improve the cooling effect, heat emission fins9 are formed outside the case 1 in parallel with the driving shaft. Theair fan for cooling the air is preferably mounted to the driving pulley5.

The case head 30 serves as a flow channel for circulating the gas whichis supplied from the gas guide member 20 into the crank chamber 70 to betaken into the cylinder bores 61 and selectively discharging the gascompressed in the cylinder to the outside. Therefore, the gas flowchannel formed to the case head 30 serves an important role to achievethe object of the present invention. The gas flow channel of the casehead 30 is mainly formed between the guide groove 32 and the plategroove 34 with separated intake and exhaust channels.

The gas intake channel is, as shown in FIG. 3, formed of a first intakeport 37 a and a second intake port 38 a. The first intake port 37 a(FIG. 3B) is communicated with the guide groove 32 at a side to beconnected to the intake manifold 23 of the gas guide member and isconnected to a side surface of the case head 30 by a side port 37 b atthe other side. The side port 37 b is formed at a lower part of themounting groove 33 to be secured by the thrust bearing 18, so that theintake gas is prevented from passing by the thrust bearing 18. Thesecond intake port 38 a (FIG. 3C) is communicated with the plate groove34 at a side to face the cylinder bores 61 and is communicated with thedriving shaft 10 by a side port 38 b to face the axial port 16 of thedriving shaft at the other side. According to the above gas intakechannels, the gas supplied from the intake manifold 23 of the gas guidemember is introduced into the first intake port 37 a and the side port37 b, and is taken into the cylinder bores 61 via the side port 38 b andthe second intake port 38 a after passing through the crank chamber 70,the block chamber 63 and the shaft chamber 15.

The gas exhaust channel is, as shown in FIG. 3A, formed of a firstexhaust port 41, a second exhaust port 42 and third exhaust ports 43 aand 43 b, wherein the first to third exhaust ports 41, 42, 43 a and 43 bpenetrate the case head 30 to be connected to the exhaust manifold 24 ofthe gas guide member. The respective exhaust ports 41, 42, 43 a and 43 bincorporate a check valve 46 for preventing the compression gas fromflowing back into the cylinder bores 61 through the exhaust manifold 24.In FIG. 3A, reference number 44 represents a drain port for discharginglubrication oil which is taken in into the shaft chamber 15 to the crankchamber, and 45 represents bolt holes for fixing the valve plate 50 tothe case head 30.

FIG. 4 shows the tension ring 40 to be inserted in the ring groove 35 ofthe case head. The ring groove 35 incorporates a circular plate typering-shaped plate spring 49 (FIG. 1) for applying elasticity to thetension ring by the plate spring 49 for the tension ring 40 to press thevalve plate 50 against the cylinder head 13.

The tension ring 40 inserted into the ring groove 35 prevents theleakage of gas to be compressed during the gas compression stroke andthe introduction of impurities such as the cooling oil into the cylinderbores 61. The tension ring 40 is divided into an intake section 40 a andfirst to third exhaust sections 40 b, 40 c and 40 d, for preventingleakage of the gas to be compressed during the gas compression strokefrom one section to another section. The tension ring 40 is formed of amaterial having heat-resistance and elasticity like a heat-resistantrubber or urethane.

The valve plate member 50 is, as shown in FIG. 1 and FIG. 5, formed inthe shape of a circular plate type ring, and has a top surface incontact with an outer surface of the cylinder head to carry out asliding motion in association with the rotation of the cylinder head.The top surface of the valve plate member 50 is formed with a singleintake valve groove 51 and separated first to third exhaust valvegrooves 52, 53 and 54 on the top surface in the shape of an arc, whereinwidths of the intake valve groove 51 and the third exhaust valve groove54 are equal to or larger than a diameter of the gas holes 14 of thecylinder head, and widths of the first and second exhaust valve grooves52 and 53 are smaller than the diameter of the gas holes.

Now, positions of the valve grooves 51, 52, 53 and 54 of the valve platemember 50 will be explained in more detail with reference to FIG. 5 andFIG. 9A. The intake valve groove 51 is positioned within a section of180° of a circumference corresponding to an intake stroke section inwhich a specific piston moves from top dead center to bottom deadcenter, and the three gas exhaust valve grooves 52, 53 and 54 are formedin the remaining 180° section of the circumference corresponding to acompression stroke section in which the piston moves from bottom deadcenter to top dead center.

A radius VR of a circumference which connects each center line of thevalve grooves 51, 52, 53 and 54 is equal to a radius HR of acircumference which connects six center lines of the gas holes 14 of thecylinder head, so that the respective gas holes 14 pass through thevalve grooves 51, 52, 53 and 54 in sequence.

The gas intake valve groove 51 and the gas exhaust valve grooves 52, 53and 54 of the valve plate member are formed apart from one another by atleast a certain distance, that is, a length VL of a partition wall,wherein it is important to keep the distance larger than a diameter ofthe gas holes 14 of the cylinder head.

It is also important to form the respective lengths of the first tothird exhaust valve grooves 52 to 54 smaller than a distance between thegas holes 14 so as not to position more than one gas hole 14 in one ofthe exhaust valve grooves 52 to 54.

The intake and first and second exhaust valve grooves 51 to 53 arerespectively formed with valve holes 51 a to 53 a penetrating the valveplate 50, and the third exhaust valve groove 54 is formed with two valveholes 54 a and 54 b.

The intake valve hole 51 a which penetrates the valve plate is connectedto the second intake port 38 a on the bottom surface of the case head,the first exhaust valve hole 52 a is connected to the first exhaust port41, the second exhaust valve hole 53 a is connected to the secondexhaust port 42, and the third exhaust valve holes 54 a and 54 b areconnected to the third exhaust ports 43 a and 43 b. Therefore, the gas,which is introduced from the second intake port 38 a of the case headinto the intake valve groove 51 of the valve plate, is taken into therespective cylinder bores 61 via the gas holes 14 of the cylinder headrotating to be gradually compressed by the rotation of the cylinderblock 60 and selectively discharged to the exhaust valve grooves 52 to54 of the valve plate which meets the gas holes 14 according to thepressure of the compression tank. The compression gas introduced intothe respective exhaust valve grooves 52 to 54 is respectively dischargedvia the exhaust valve holes 52 a, 53 a, 54 a and 54 b and the exhaustports 41, 42, 43 a and 43 b to the exhaust manifold 24 of the gas guidemember.

As shown in FIG. 1, the cylinder block 60 is formed in the shape of acylinder on the whole, and it is formed with the block chamber 63 in theaxial center of the cylinder block. Also, the cylinder block 60 isformed with the six cylinder bores 61 of an equal diameter radiallyadjacent to the block chamber 63 in a longitudinal direction of andparallel with the driving shaft.

The cylinder block 60 is coupled and sealed with the cylinder head 13 ofthe driving shaft by bolts at a sectional surface, and the respectivelycylinder bores 61 are slidably inserted with six pistons 64. The blockchamber 63 is fixed with a coupling 65 by a bolt for coupling to auniversal coupling 66, which is explained hereinafter. The cylinderblock 60 is formed with spiral heat emission fins 67 on an outerperipheral surface, wherein the heat emission fins 67 accelerate thecirculation of the gas that flows through the space in the crankchamber, when the cylinder block 60 rotates. The heat emission fins 67may also be formed in a plurality of circles apart from one another by auniform interval in addition to the spiral shape. The block chamber 63is formed in the shape of a cylinder of two stages having differentdiameters for preventing the dispersed lubrication oil from flowingtoward the chamber 15.

The cylinder block 60 is connected to the swivel plate member 80 by theuniversal coupling 66 along central lines of rotations axis ofrespective parts, wherein the swivel plate member will be describedbelow. The universal coupling 66 is formed of a driving joint 68 and adriven joint 69, wherein the driving joint and the driven joint areconnected to each other by a cross shaft 71 at each fork-type arm part.The driving joint 68 of the universal coupling is hollow and has acylindrical spline at an outer periphery to be spline-coupled with thecoupling 65 that is fixed to the cylinder block 60, and the driven joint69 is formed with a flange at an end to be coupled with the swivel platemember by bolts.

The swivel plate member 80 is axially coupled with a driven shaft 7 byway of a taper roller bearing 81 in the center of its rotation shaft,wherein the driven shaft 7 is fixed to a slant surface 6 of the case endplate 3 in the center of its rotation shaft. The swivel plate member 80is formed with 6 pairs of brackets 82 on an outer periphery thereof,each pair of brackets to be coupled to one end of a piston rod 73, andit has a shoulder on the opposite surface to the brackets 82 to mount athrust bearing 83 in a space between the slant surface 6 and the swivelplate member 80.

The six pistons 64 respectively inserted into the cylinder bores 61 arerotatably coupled with the brackets 82 of the swivel plate member by thepiston rods 73. The piston rods 73 coupled between the respectivepistons 64 and the swivel plate member 80 may be selected from auniversal-coupling or a two-fold type crank. A piston rod 73 in theshape of the two-fold type crank is formed of a first rod 74 and asecond rod 76. The first rod 74 is connected to the piston 64 by aconnection pin 75 at one end and is formed with a fork arm at the otherend. The second rod 76 is connected to the bracket 82 of the swivelplate member by a connection pin 77 at one end, and it is formed with acoupling hole at the other end. The first rod 74 and the second rod 76are coupled by inserting a connection pin 78 through the fork arm of thefirst rod 74 and the coupling hole of the second rod 76 such that therods 74 and 76 are rotatably fixed by the connection pin 78. At eachconnection point, bearings are coupled with outside the connection pins.The piston rod 73 in the shape of a universal coupling, which is notshown in the drawings, has a similar coupling structure as the pistonrod in the shape of the twofold type crank, wherein both rods arerotatably fixed by the cross shaft.

FIGS. 6A through 6C and FIG. 7 show further embodiments of the presentinvention.

In the embodiments of the present invention, main parts which determinethe driving mechanism of the present invention have the same functions,and various modifications are made with respect to the structure of thegas intake channel, the connection structure between the cylinder block60 and the swivel plate member 80, and the cooling structure for coolingthe gas compressor.

The embodiment of the present invention as shown in FIG. 6 is basicallyequivalent to the embodiment of the present invention as shown in FIG.1, except for the connection structure between the cylinder block 60 andthe swivel plate element 80. In FIG. 6A, the piston rods 73 whichconnect the respective pistons 64 to the swivel plate member 80 areformed in the two-fold type crank shape as in the embodiment of FIG. 1,whereas the cylinder block 60 and the swivel plate member 80 areconnected to each other not by the universal coupling but by a springcoupling 72. Therefore, the power of the driving shaft 10 is transmittedto the swivel plate member 80 not by the universal coupling but by thepiston rods 73.

The spring coupling 72 connects the cylinder block 60 to the swivelplate member 80 with two parallel springs 72 a and 72 b as shown in FIG.6B. The cylinder block 60 is formed with two block rings 68 a, 68 bdiagonally across from each other with respect to the springs 72 a and72 b, and the swivel plate member 80 is formed with first and secondswivel plate rings 69 a, 69 b also diagonally across from each other, sothat the first spring 72 a is coupled with the first block ring 68 a andthe first swivel plate ring 69 a, and the second spring 72 b is coupledwith the second block ring 68 b and the second swivel plate ring 69 b.According to the coupling structure of the spring coupling 72, thecylinder block 60 and the swivel plate member 80 apply attraction forceto each other in the same direction as the rotation of the cylinderblock 60 and the piston rods 73, as shown by an arrow of FIG. 6B. Theattraction force compensates the reaction force which is generated bythe swivel plate member 80 when the cylinder block 60 begins to rotate,wherein the springs 72 a and 72 b of the spring coupling 72 respectivelyhave a coefficient of elasticity determined in consideration of thereaction force of the swivel plate member 80. In the block rings 68 aand 68 b and the swivel plate rings 69 a and 69 b, holder parts in whichfixing pins for fixing the respective rings are connected to the springsare connected by ball joints in consideration that the swivel platemember 80 carries out a swing motion when the gas compressor operates.

FIG. 6C shows a spring coupling 72 according to another embodiment ofthe present invention. In FIG. 6C, the spring coupling 72 is formed of acylinder 72 c of which ends are respectively coupled with a block flange68 c and a swivel plate flange 69 c, and the block flange 68 c and theswivel plate flange 69 c are respectively fixed to the cylinder block 60and the swivel plate member 80 by bolts. When coupling the cylindricalspring 72 c, the cylindrical spring 72 is initially distorted by apredetermined amount in the same direction as the rotation of thecylinder block 60 and the piston rods 73. Therefore, the distortionstress of the spring 72 compensates the reaction stress generated by theswivel plate member 80 when the cylinder block 60 begins to rotate.

Comparing the gas compressor as shown in FIG. 7 with that of FIG. 1, thestructure of the gas intake channels, the connection structure betweenthe cylinder block 60 and the swivel plate member 80, and the coolingstructure for cooling the gas compressor are modified. Now, the mainparts of the modified embodiment as above will be explained in detail.

The differences of the embodiment of FIG. 7 from that of FIG. 1 are asfollows.

First, the driving shaft 10 is formed with a through hole 15 a whichcompletely penetrates the inside of the driving shaft, axially. In thecase that the through hole 15 a is formed in the driving shaft, theaxial port 16 of FIG. 1 is omitted. The through hole 15 a serves as achannel for discharging lubrication oil mist which is generated in thecrank chamber 70 to the outside and as a cooling tube channel forintroducing atmospheric air into the crank chamber 70, when the gascompressor using the lubrication oil is in operation.

Second, the intake port 39 of the case head member 30 is formedpenetrating a portion between the guide groove 32 and the plate groove34. Comparing the two embodiments of FIG. 1 and FIG. 7, the first andsecond intake ports 37 a and 38 a (shown in FIG. 3A), the side ports 37b and 38 b, and the additional drain port 44 as shown in FIG. 3A are notformed in the case head 30 of FIG. 7. The direct penetration of the gasintake port 39 is to supply the gas introduced from the gas guide memberto the inside of the cylinder bores 61 without circulating it inside thecrank chamber 70.

Third, the cylinder block 60 and the swivel plate member 80 are coupledwith two shafts by a bevel gear 86, improving the assemblingperformance.

Fourth, the pistons 64 and the swivel plate member 80 are connected topiston rods 73 in the shape of a two-fold type extension rod. Byseparately forming the piston rods in two parts in the shape of male andfemale bolts, a stroke clearance of the pistons may be finely controlledwhen assembling the pistons by using lock nuts 84. If the extension typerod is used for the piston rods 73 as shown in FIG. 7, a plate spring 85is mounted outside the respective piston rods and on the outerperipheral surface of the swivel plate member 80. The plate spring 85serves to compensate the centrifugal force applied to the pistons 64when the cylinder block 60 rotates with the swivel plate member 80.

As shown in FIG. 7, if the extension rod type piston rods 73 areemployed, it is preferable to operate the compressor by pouring thelubrication oil into a bottom part of the crank chamber 70. If thelubrication oil is poured into the bottom part of the crank chamber 70for operating the compressor, the lubrication oil is dispersed into thechambers when the compressor is operating, thereby cooling friction heatthat is generated between balls 87 of the extension rods and ball joints88.

Fifth, the cylindrical case 1 is attached with a cooling case 90surrounding the outside of the cylindrical case. Therefore, a coolingchamber 8 is formed between the case 1 and the cooling case 90. Thecooling case 90 is formed with a coolant intake hole 90 a and a coolantdischarge hole 90 b, and heat emission fins 9 a are spirally formedoutside the case 1 for channeling the coolant circulating in the coolingchamber 8 to the outside via the discharge hole 90 b after it circulatesover the outer peripheral surface of the case 1. The heat emission fins9 and 9 a are arranged spirally as shown in FIG. 7 for a liquid coolingsystem, and in parallel to the driving shaft 10 as shown in FIG. 1 foran air cooling system.

Finally, a plurality of blades 89 is formed on an outer peripheralsurface of the swivel plate member 80 in the gas compressor as shown inFIG. 7. The blades 89 serve to disperse the lubrication oil from thebottom of the crank chamber 70 to the inside of the chamber in thecompressor that uses the lubrication oil.

The above preferred embodiments and modifications thereof are explainedwith respect to the main parts required for operation and the couplingrelationship therebetween. Explanation of the other parts such as casesealing parts, sliding balls, piston rods and the like of whichstructures are similar to those of general mechanical equipment will beomitted.

The axial coupling structure of the cylinder block 60 and the swivelplate member 80, the piston rods 73, and the cooling structure explainedwith reference to the preferred embodiments and modifications of thepresent invention are not limitedly used alone as in the respectivecorresponding embodiments, but may be used in combination selectivelyaccording to the usage of the gas compressor.

Now, the operation and the operational characteristics of the gascompressors according to the preferred embodiments and modifications ofthe present invention as described above will be explained in detailwith reference to FIG. 8 to FIG. 10.

As shown in FIG. 8, the rotation force generated from the external powersupply such as a motor (not shown) is transmitted to the pulley 5 via apower transmission element such as a belt (not shown). As the drivingshaft 10 rotates by the rotation force transmitted to the pulley 5, thecylinder head 13 and the cylinder block 60 rotate together with thedriving shaft 10. Simultaneously, the swivel plate member 80 coupledwith the cylinder block 60 by the universal coupling 73 (or the springcoupling as shown in FIG. 6, or the bevel gear as shown in FIG. 7)rotates with respect to the driven shaft 7. As the swivel plate 80swings in the direction inclined toward the driving shaft, therespective piston rods 73 coupled with the swivel plate member 80reciprocate rectilinearly in the driving shaft direction.

The power of the driving shaft 10 in FIGS. 6A and 6C is transmitted tothe swivel plate member 80 not by the spring coupling 72 but by thepiston rods 73. At this time, the respective springs 72 a, 72 b and 72 cof the spring coupling 72 are applied with attraction force ordistortion stress, so that the reaction force which is generated by theswivel plate member 80 when the cylinder block 60 begins to rotate, iscompensated by the attraction force or the distortion stress.

The motion of the compressor parts as above is carried outsimultaneously with the input of the power, and the six pistons 64 carryout the exhaust stroke selectively while rotating together with thecylinder block 60 and simultaneously reciprocate respectively. In astroke distance of the pistons 64, connection points between the swivelplate member 80 and the piston rods 73 are equal to a distance that theswivel plate moves in the driving shaft direction in the swing motion,which may be represented by 2R sin(K0), wherein R represents a distancefrom a center of the driven shaft 7 to the connection points between theswivel plate member 80 and the piston rods 73, and K0 represent aninclination angle of the driving shaft 10 and the driven shaft 7.

The flow of the gas in the compressor will be explained below.

First, the gas passes through the external filter (not shown) and isintroduced into the intake tube 21 of the gas guide, at which point itcirculates in the crank chamber 70 and is introduced into the cylinderbores 61 in the embodiments of FIG. 1 and FIG. 6, while it is directlyintroduced into the cylinder bores 61 in the embodiment of FIG. 7. Asshown in FIG. 1, the gas circulation path for the case in which the gasis introduced after circulation is explained below in detail.

The gas introduced via the intake tube 21 of the gas guide passesthrough the first intake port 37 a of the case head, the side port 37 b,the crank chamber 70, the block chamber 63, the shaft chamber 15, theaxial port 16, the side port 38 b, the intake valve hole 51 a of thevalve plate and the gas holes 14 in sequence and it is then introducedinto the cylinder bores 61. An object of the introduction of the intakegas after circulation in the crank chamber 70 instead of directlyintroducing it into the cylinder bores 61 is to buffer noise caused byresidual pressure that may remain in the cylinder bores 61 after thecompression and exhaust strokes, as the gas introduced into the crankchamber 70 at a low pressure induces an explosion, which is generated atan instant that gasses of different pressure are mixed, and in thesealed space, that is, in the crank chamber 70, the noise due to themixing of gases having different pressures is muffled.

The gas introduced into the cylinder bores 61 is compressed while thecylinder block 60 and the pistons 64 rotate, and is dischargedselectively according to a pressure of the compression tank via therespective exhaust valve holes 52 a, 53 a, 54 a and 54 b of the valveplate and the respective exhaust ports 41, 42, 43 a and 43 b of the casehead at an instant when the gas holes 14 of the cylinder headrespectively meet the first to third exhaust valve grooves 52 to 54 ofthe valve plate, in sequence. The discharged gases are channeled to theexhaust manifold 24 of the gas guide member and are discharged via theexhaust tube 22.

Friction heat, which is generated by friction between the respectiveparts in the operation of the compressor, may be cooled by abelow-mentioned manner.

In the case that lubrication oil is used in the embodiment of FIG. 7,the compressor is operated under a state whereby the lubrication oil isadded to the crank chamber 70 until the blades 84 of the swivel platemember are immersed in the lubrication oil. In the above compressor, theblades 84 of the rotating swivel plate scatter the lubrication oil ontoinner walls of the crank chamber 70 and simultaneously the heat emissionfins 67 of the cylinder block stir the lubrication oil remaining in thesump. Therefore, the scattered lubrication oil is supplied to each ofthe operating parts and simultaneously cools the friction heat generatedby the friction of the parts. At this time, if the lubrication oil ispartially atomized and becomes an oil-vapor state, the oil-vapor isdischarged to the outside of the compressor via the through hole 15 a ofthe driving shaft. Further, the compression heat generated in thecylinder bores 61 and emitted toward the block chamber 63 is also cooledby the vortex flow of the gas formed in the crank chamber 70.

In case that the compressor is operated without using the lubricationoil, the compressor is combined in a power transmission structure inwhich the friction parts may be minimized. In the case of thecompressor, the gas introduced into the crank chamber 70 circulateswhile forming the vortex flow in the chamber, so that the circulatinggas itself serves as a cooling medium.

Even though the cooling operation in the crank chamber 70 is explainedin the above, the compressor according to the present invention coolsthe outside of the case 1 as well as the inside of the compressor in anair or liquid cooling manner. The air-cooling is carried out by an airfan, with radial heat emission fins 31 formed on the case head radiallywith respect to the driving shaft and heat emission fins 9 formed on thecase in parallel with the driving shaft. That is, the air fan is mountedto the motor (not shown) which is positioned in front of the drivingshaft for generating wind toward the case 1, so that the wind flowsalong the heat emission fins 31 and 9 outside the case and cools theoutside of the compressor. The liquid cooling is carried out by coolantsupplied toward the intake hole 90 a of a cooling chamber 8 between thecase 1 and the cooling case 90, which flows along the spiral heatemission fins 9 a and is discharged via the discharge hole 90 b aftercirculating over the outer peripheral surface of the case 1.

The pistons 64 which rotate together with the cylinder block 60 areapplied with centrifugal force in a direction such that a radiusincreases with respect to the driving shaft 10. In order to compensatethe centrifugal force applied to the pistons, as shown in FIG. 7, theplate spring 85 is mounted to the piston rods 73, thereby compensatingthe centrifugal force applied to the pistons 64 that are in motion.

Now, the compression and exhaust strokes carried out in the compressoraccording to the present invention will be described in more detail.FIGS. 9A and 9B show intake, compression and exhaust characteristicswhen the respective valve grooves 51 to 54 meet the respective gas holes14 of the cylinder head 13 in the process of reciprocation of thepistons 64.

In FIG. 9A, reference symbol T represents a position in which the piston64 is located at top dead center, and B represents a position in whichthe piston 64 is located at bottom dead center. Reference symbol Krepresents an angle that the gas holes 14 of the cylinder head arerotated around the valve plate 50. In FIG. 9A, if the gas holes 14rotate in the counterclockwise direction around the valve plate 50, asection in which K=0°˜180° corresponds to the intake stroke section ofthe piston travel and a section in which K=180°˜360° corresponds to thecompression and exhaust stroke section of the piston travel.

In section S1 in which the gas holes 14 rotate from top dead center T toa position K1 immediately before meeting the intake valve groove 51, thepartial compressed gas which is not exhausted but remains in thecylinder bores 61 is expanded. In section S2, the gas holes 14 passthrough the intake valve groove 51 for intake of the gas. From aposition K2 to the bottom dead center B position at K3, the valve plate50 closes the gas holes 14 for preparing for compression. R1 representsa section where the gas holes 14 move to a position K4 as they are beingclosed, wherein the intake gas is primarily compressed. E1 is a sectionwhere the gas holes 14 pass the first exhaust valve groove 52 andprimarily exhaust the primarily compressed gas. R2 represents a sectionin which the gas holes 14 move to a position K6 of being closed again,wherein the primarily compressed gas is secondarily compressed. E2 is asection in which the gas holes 14 pass the second exhaust valve groove53 and secondarily exhaust the secondarily compressed gas. R3 is asection in which the gas holes 14 are closed again and they move to aposition K8, wherein a tertiary compression is carried out on thesecondarily compressed gas. E3 is a section in which the gas holes 14pass the third exhaust valve groove 54 for carrying out a third exhaustof the tertiarily compressed gas. From K9 to K10, the valve plate 50closes the gas holes 14 again, and the pistons compress the gas up totop dead center T position and prepare for the next intake stroke.

The significant characteristics of the present invention lie on thecompression and exhaust strokes. In the exhaust stroke, the exhaustvalve grooves act as the exhaust stroke section when the pressure of theexhaust manifold 24 is low but as the compression stroke section whenthe pressure of the exhaust manifold 24 is high. That is, even thoughthe gas holes 14 meet the first to third exhaust valve grooves 52 to 54in the exhaust sections E1 to E3, the compressed gas may be dischargedvia the exhaust valve grooves 52 to 54 only when the pressure of thecompressed gas is higher than the internal pressure of the exhaustmanifold 24. If the pressure of the gas compressed in the cylinder bores61 is lower than the internal pressure of the exhaust manifold 24, thecheck valves 46 mounted in the respective exhaust ports become closed,so that backflow from the exhaust manifold 24 to the cylinder bores 61may be prevented and the above sections serve as the compression strokesections instead of the exhaust stroke sections.

FIG. 9B shows a pressure P of the gas that may be obtained in thecylinder bores 61 for each rotation angle in the case that all the checkvalves are closed and one of the cylinder bores 61 rotates by one cyclealong the valve plate 50.

At this time, the pressure loss due to the check valves and the exhaustchannel are ignored for the sake of convenience of explanation, and itis assumed that the pressure in the exhaust tube 22 is equal to that inthe compression tank. If the pressure in the intake tube 21 is P00 and apressure when the gas is taken into the cylinder bores 61 and therotation angle becomes K1 is P0, it is assumed that P00>P0 due to theair friction loss generated in the process of the intake. Then, Ptrrepresents a rated pressure in the compression tank, and Pmax representsan available maximum pressure that may be obtained in the cylinder bores61 while all of the check valves are closed. When actually designing acompressor, the rated pressure Ptr in the compression tank is set asshown in FIG. 9B and the stroke distance of the pistons are controlledto keep the available maximum pressure Pmax in the cylinder bores 61higher than the set rated pressure Ptr. As the rotation angle K of thegas holes 14 becomes K4, the gas pressure in the cylinder bores 61 at apoint 4 becomes P4. Subsequently, if the rotation angle of the gas holes14 becomes K5, K6, K7, K8, and K9 in sequence, the gas pressure becomesP5, P6, P7, P8 and P9 in sequence. Therefore, when designing thecompressor according to the present invention, the rated pressure Ptr ofthe compression tank becomes the pressure between P8 and P9 that may beobtained when the gas holes 14 are positioned to the final exhaustsection E3. As shown in the dotted line of FIG. 9B, the position K1 isset to equalize the pressure of the gas remaining in the cylinder bores61 after the compression and exhaust strokes with the intake pressureP0.

In the case that the compressor according to the present invention iscontinuously operated on the basis of the intake and compression/exhauststroke characteristics and the pressure change in the compressor, thecompression characteristics of the gas will be explained below withreference to FIG. 10.

FIGS. 10A to 10C show the compression characteristics of the compressorwhich may be obtained by a single cylinder bore 61, wherein FIG. 10Ashows the compression characteristics of the compressor according to thepresent invention, FIG. 10B shows the compression characteristics of aprior art reciprocating compressor, and FIG. 10C shows the compressioncharacteristics of a prior art rotary slant shaft type compressor.

In the figures, the horizontal axis represents the number ofreciprocation stroke of the pistons, which is equal to a rotation numberN of the cylinder block 60. First, the compression characteristics asshown in FIG. 10A will be explained in detail. In the case that thepressure Pt of the compression tank as shown by points D corresponds tothe pressures between point P4 and point P5 which represent the rotationpositions K4 and K5 of FIG. 9A, the pressure in the cylinder bores 61changes from point 3, point 4, point 4D, point 5, point 6, point 7,point 8 and point 9 in sequence. That is, the compression is carried outfrom the initial pressure P0 to the point 4 and the compression ispreceded by the point 4D under the state that the check valve 46 in thefirst exhaust port 41 is closed. However, passing the point 4D, thecheck valve 46 in the first exhaust port 41 is opened and the samepressure is continued to the point 5. Further, from the point 5 to thepoint 6, the gas holes 14 passing through the section R2 of FIG. 9A areclosed, thereby proceeding with the compression. Next, at the point 6where the secondary exhaust section E2 begins, the pressure in thecylinder bores 61 becomes higher than the pressure of the point D whichrepresents the pressure of the compression tank, so that the check valve46 in the secondary exhaust valve 42 opens and the pressure istemporarily decreased by the point 7 where the secondary exhaustprocedure E2 is finished. Next, from the point 7 to the point 8, the gasholes 14 are closed again and the section R3 of FIG. 9A is passed,thereby proceeding with the compression again. At the point 8 where thethird exhaust section E3 begins, the pressure in the cylinder bores 61becomes higher than the pressure in the compression tank, so that thecheck valve 46 in the third exhaust port 43 is opened and the pressuredecreases to the pressure of the point D by the point 9 where the thirdexhaust procedure E2 finishes.

In the above procedure, if the pressure P of the cylinder bore is lowerthan the pressure Pt of the compression tank as shown by the point D,the check valve 46 remains closed, while it remains open if the pressureP of the cylinder bore is higher than the pressure pt of the compressiontank. The reference symbol D in FIG. 10A represents a position where thecheck valve 46 opens. Therefore, as the number of rotations of thecompressor becomes larger, the pressure D of the compression tankbecomes higher and the pressure discharged from the cylinder bore 61 tothe compression tank is changed from the point 4 to the point 9 as shownby a dotted line of FIG. 10A.

On the other hand, if the compressor is operated under a state such thatthe auxiliary exhaust tube 26 is connected to the auxiliary intake tube25 while the compressor is continuously operating, that is, the insideof the cylinder bore 61 is not applied with any compression load, whichis the loadless operation state, the compression characteristics of sucha compressor are represented as in the right part of FIG. 10A. In thiscase, a certain pressure loss is generated during intake of outside gashaving the pressure P00 into the cylinder bores 61. Considering such apressure loss, the pressure in the cylinder bores 61 becomes P0, andthis pressure is to be the pressure at the point 3. If the compressionstroke is proceeded under the loadless operation state, a pressure curveof the cylinder bores 61 has partial compression sections from the point3 to the point 4, from the point 5 to the point 6, and from the point 7to the point 8. However, the pressure in the cylinder bores finallybecomes equal to the pressure P00 of the outside gas to be inhaled,since the compression and exhaust strokes are carried out while all thecheck valves 46 are opened.

Now, the compression characteristics of the prior reciprocation typecompressor and the prior art slant shaft type compressor will bedescribed in more detail for a comparison with the compressioncharacteristics of the compressor according to the present invention.

Referring to FIG. 10B, the prior art reciprocation type compressor isinitiated to operate under the state that the pressure at the point Dwhich represents the pressure Pt of the compression tank is lower thanthe rated pressure Ptr, the pressure of gas to be compressed in thecylinder as shown by a point B becomes higher than the pressure of thecompression tank before the piston reaches top dead center, so that theexhaust valve is opened immediately and the gas is discharged. That is,if the pressure P of a cylinder chamber is lower than the pressure D ofthe compression tank in the gas compression stroke, the compressioncontinues. If the pressures become equal, an exhaust valve opens forcarrying out the exhaust. Therefore, as the number of compressionsincreases, that is, as the pressure of the compression tank becomeshigher, the position of the point B where the compression stroke isfinished for each rotation becomes higher. Further, in the case ofloadless operation, the exhaust valve is opened at a time point H whenthe pressure of the exhaust tube becomes P00, as shown in the right partof the compression characteristics curve of FIG. 10B.

In the case of the prior art rotary slant shaft compressor as shown inFIG. 10C, the gas in the cylinder chamber is always compressed up to thepoint B, it is exhausted when the point B is higher than the point Dwhich represents the pressure of the compression tank, and it continuescompression if lower. Even in the case of loadless operation, the gas iscompressed up to the point H and then the pressure of the gas isimmediately lowered to the pressure P00 of the outside gas to beinhaled.

As shown in FIG. 10, if the operation of the compressor is changed toloadless operation at the point where the pressure of the compressiontank reaches the rated pressure Ptr during the operation of thecompressor, the energy efficiency may be improved. Therefore, a totalload amount of the compressor that is required for the gas compressionis the sum of polygonal areas formed by the point 3 to the point 9 orthe points A, B, C and F for each revolution. The total load amount ofthe compressor is proportional to the total energy amount that isrequired for driving the compressor.

According to the compressor of the present invention, the total energyconsumption required for compression is similar to that of the prior artreciprocating compressor as show in FIG. 10B, but much smaller than thatof the prior art slant shaft type compressor as shown in FIG. 10C.

Therefore, the compressor according to the present invention has higherenergy efficiency in comparison with the prior art slant shaft typecompressor. In particular, even in the case of loadless operation, thecompressor of the present invention exhibits energy consumption that isnoticeably smaller than the prior art slant shaft type compressor.

On the other hand, noise, which is generated in the compressor, becomeslarger as a difference of pressure between the inside of the cylinderand the inside of the compression tank becomes larger. The prior artslant shaft type compressor, as shown in FIG. 10C, exhibits a largepressure difference between the point B and the point D, while thecompressor of the present invention, as shown in FIG. 10A, exhibits asmall pressure difference between the point D and the sections from thepoint 5 to the point 6 and from the point 7 to the point 8. This resultshows that there is a very small pressure difference between thecompressed gas in the cylinder and the compressed gas in the compressiontank, so that the explosion generated when gasses of different pressuresmix is very slight. Therefore, the gas compressors according to theembodiments as shown in FIG. 1, FIG. 6 and FIG. 7 have an advantage inthat they exhibit little noise.

Considering that the compression load applied to the cylinder bores 61are equal to the axial force load applied to the driving shaft 10, thecompressor of the present invention, as shown in FIG. 10, exhibits avery small change of the compression load per unit time period incomparison with the prior art slant shaft type compressor. Therefore,according to the compressor of the present invention, it is possible toincrease the durability of the bearings which support the swivel plate80 which receives the influence of the variable load directly, as wellas the bearings connected to the driving shaft 10.

As described hereinabove, the compressor according to the presentinvention may carry out the exhaust stroke selectively according to thepressure of the compression tank, and it may be designed in thestructure such that the gas to be introduced into the cylinder bores istaken into the cylinder bores directly or after circulation toward thecrank chamber for obtaining the following effects.

First, the compressor may be operated quietly by reducing the noisesource aerodynamically.

Second, the energy efficiency is maximized by minimizing the powerrequired for the gas compression.

Third, the durability of the compressor is improved by reducing thechange of the compression load per unit time period.

Fourth, loadless operation of the compressor is possible with highefficiency.

Fifth, the gas compression efficiency and the durability of thecompressor may be improved by circulating coolant around the compressorand emitting the heat generated by the mechanical friction and the aircompression by using cooling lubrication oil.

Sixth, the heat generation may be restrained and the lifespan of thecompressor may be extended by compensating the centrifugal forcegenerated by the rotation of the pistons and reducing the relativefrictional force generated on contact surfaces in the cylinder, andSeventh, the assembling productivity of the compressor may be improvedby simplifying the assembling of the pistons by forming the piston rodsas the two-fold type.

It will be apparent to those skilled in the art that variousmodifications and variations can be made to the device of the presentinvention without departing from the spirit and scope of the invention.The present invention covers the modifications and variations of thisinvention provided they come within the scope of the appended claims andtheir equivalents.

What is claimed is:
 1. A rotary slant shaft type gas compressorcomprising: a driving shaft integrally formed with a cylinder headperpendicular to the driving shaft, the cylinder head being formed witha plurality of gas holes on a concentric circle at uniform intervals; agas guide member formed with an intake manifold for intake of gas fromthe outside and an exhaust manifold for discharging gas compressed incylinder bores to the outside; a case head member for rotatablysupporting the driving shaft formed with at least one intake port forsupplying the gas taken in through the intake manifold to the inside ofthe cylinder bores, and two or more exhaust ports for discharging thegas compressed in the cylinder bores to the exhaust manifold; a valveplate member fixed on an inner surface of the case head member tocontact an outer surface of the cylinder head, and formed with a gasintake valve groove and at least two gas exhaust valve grooves on aperiphery on which the gas holes move, the gas intake valve groovesupplying the gas taken in through the intake port to the inside of thecylinder bores and the gas exhaust valve grooves discharging the gascompressed in the cylinder bores to the exhaust ports; a cylinder blockformed with a plurality of cylinder bores in parallel with the drivingshaft and having a surface integrally coupled with the cylinder head,and having an opposite surface slidably inserted with pistons inrespective cylinder bores for compressing the intake gas in therespective cylinder bores; a swivel plate member connected to a centerpart of the cylinder block with a coupling, and connected to a pluralityof pistons via piston rods, for converting a rotation force transmittedfrom the driving shaft to rectilinear reciprocation motion to betransmitted to the pistons; a case end plate formed with a slant surfacefor supporting the swivel plate member; and a case coupled with the casehead member and the case end plate for incorporating the cylinder blockand the swivel plate member.
 2. A rotary slant shaft type gas compressoras claimed in claim 1, wherein the respective exhaust ports of the casehead member incorporate respective check valves.
 3. A rotary slant shafttype gas compressor as claimed in claim 2, wherein the gas intake valvegroove of the valve plate member is formed within a section of 180° of acircumference corresponding to an intake stroke section in which aspecific piston moves from top dead center to bottom dead center, andthe gas exhaust valve grooves are formed in a remaining 180° section ofthe circumference corresponding to a compression stroke section in whichthe piston moves from bottom dead center to top dead center.
 4. A rotaryslant shaft type gas compressor as claimed in claim 3, wherein the gasintake valve groove and the gas exhaust valve grooves of the valve platemember are formed apart from one another by at least a certain distance,that is, a length (VL) of a partition wall, which is larger than adiameter of the gas holes of the cylinder head.
 5. A rotary slant shafttype gas compressor as claimed in claim 4, wherein each length of thegas exhaust valve grooves of the valve plate member is shorter than adistance between the gas holes.
 6. A rotary slant shaft type gascompressor as claimed in claim 5, wherein each width of the gas valvegrooves of the valve plate member is formed the same as or larger thanthe diameter of the gas holes.
 7. A rotary slant shaft type gascompressor as claimed in claim 6, wherein at least one of the gasexhaust valve grooves of the valve plate member is formed with a groovewidth smaller than the diameter of the gas holes.
 8. A rotary slantshaft type gas compressor as claimed in claim 5, wherein the case headmember and the driving shaft are formed with a circulation circuit forintroducing the gas introduced from the intake manifold to the cylinderbores via a sealed crank chamber formed inside the case.
 9. A rotaryslant shaft type gas compressor as claimed in claim 8, wherein thecirculation circuit is formed with at least one intake channelcommunicated from the intake manifold of the case head member to thecrank chamber and at least one sub-intake channel communicated from thecrank chamber to the cylinder bores via a cylinder block chamber, andthe driving shaft is partially hollow in an axial direction of thedriving shaft, wherein the hollow part is formed with at least one axialport perpendicular to the driving shaft.
 10. A rotary slant shaft typegas compressor as claimed in claim 9, wherein the hollow part in thedriving shaft or the cylinder block chamber is formed with at least oneliquid introduction preventing shoulder.
 11. A rotary slant shaft typegas compressor as claimed in claim 10, further comprising a tension ringand a ring-shaped plate spring inserted between the inner surface of thecase head member and the valve plate member.
 12. A rotary slant shafttype gas compressor as claimed in claim 11, wherein an outer surface ofthe case is attached with a cooling case surrounding the case, thecooling case being formed with a coolant intake hole and a coolantdischarge hole, and wherein the case is formed with spiral heat-emissionfins outside the case.
 13. A rotary slant shaft type gas compressor asclaimed in claim 12, wherein an outer peripheral surface of the swivelplate member is formed with a plurality of blades.
 14. A rotary slantshaft type gas compressor as claimed in claim 13, wherein an outersurface of the case head member is formed with heat emission finsprojecting radially with respect to the driving shaft and an outersurface of the case is formed with heat emission fins in parallel withthe driving shaft.
 15. A rotary slant shaft type gas compressor asclaimed in claim 14, wherein the gas guide member is formed with anauxiliary intake tube and an auxiliary exhaust tube that connect anintake tube communicated with the intake manifold to an exhaust tubecommunicated with the exhaust manifold.
 16. A rotary slant shaft typegas compressor as claimed in claim 1, wherein the coupling that connectsthe cylinder block to the swivel plate member comprises a universalcoupling or a bevel gear.
 17. A rotary slant shaft type gas compressoras claimed in claim 16, wherein a piston rod that connects a piston tothe swivel plate member comprises a universal coupling, a two-fold crankor an extension rod.
 18. A rotary slant shaft type gas compressor asclaimed in claim 17, wherein the extension rod is formed of a male and afemale bolt, wherein the stroke clearance of the piston is controlled byusing a lock nut coupled with an outside of the male bolt.
 19. A rotaryslant shaft type gas compressor as claimed in claim 18, wherein a platespring is mounted outside the extension rod.
 20. A rotary slant shafttype gas compressor as claimed in claim 1, wherein the coupling thatconnects the cylinder block to the swivel plate member comprises aspring coupling.
 21. A rotary slant shaft type gas compressor as claimedin claim 20, wherein the piston rod that connects the piston to theswivel plate member comprises a two-fold crank.
 22. A rotary slant shafttype gas compressor as claimed in claim 21, wherein the spring couplingis connected between the cylinder block and the swivel plate member inthe same direction as the rotation of the cylinder block and the pistonrod for applying attraction force or distortion stress.